1284719 九、發明說明： 【發明所屬之技術領域】 本發明係有關於一種輪式車 及-種具變動傳動比之防滑差迷器用防滑差速器，具體涉 【先前技術】 目前防滑差速器的結構有多種，、 為内摩擦式、超越式、電子自’工作原理可歸納 相結合的電子制動防滑差速争 叙貞及/、ABS剎車系統 六厂从 及勢壘式幾種，它們各自 存在者不同的缺點，如内摩擦式 艰1吊用的兩摩捧式# 益，又可細分為預載入式和非 ，、弋差速 沾址赴工〇 截入式，都具有價格昂貴 的缺點，而後者的價格更高， 、 ^ 考有轉向阻力較大，銓 胎磨損較快的缺點；超越式的 ι … 可罪性差，運行不平穩，結 構複雜；電子自動差速鎖同樣存 心 ^ 子在作不夠平穩及結構複 雜的缺點；與ABS剎車系統妹人沾在 …口的制動防滑系統動力消耗 較大。 該專利公開了-種勢壘式變動傳動比差速器其具有結 ㈣単、性能可靠並且不會產生負牽引力的優點，其工作 原理為行星齒輪與半㈣輪作週期性的變速比傳動，並且 行星齒輪上的速比變化週期為奇數次，故當行星齒輪與一 側半軸輪的傳動比達到最大值時，與另一侧半抽齒輪的傳 動比達到最小值，從而實現兩侧半軸齒輪上不相等的扭矩 分配；行星齒輪轉過半個速比週期後兩侧半軸齒輪上的扭 矩分配將對調過來。這種週期性的扭矩分配變化形成了兩 5163-5619-PF2 5 1284719 個勢s，如果兩側半轴輪上的扭矩之比不超過差速器的最 大扭矩分配，差速器就不能連續差動，限制了車輪滑轉。 但由於目前所用的變動傳動比差速器的速比變化週期為一 個周節’在每一個周節裏要完成一次行星輪與半軸輪的相 對加速與減速過程，相對角加速度較大，其值與速比變化 範圍成正比，與行星輪一轉中速比波動週期數的平方成正 比。車父大的相對角加速度加大了齒面間的相對曲率，降低 了齒輪對的承載能力，並容易產生雜訊。雖然通過參數優 化可以在一定程度上提高其速比波動範圍，但效果有限， 進一步提高變速比範圍將使齒面間的相對曲率迅速增大， 甚至在齒面上產生棱線。按傳統的單周節設計方法，行星 齒輪為7齒，半軸齒輪為12齒的變動傳動比差速器兩半軸 輪間由差動運動產生的最大變速比只有T:1•狀 為9齒，半軸齒輪為12齒的差速器僅能到1:1·31，對於 越野汽車仍然不足。 【發明内容】 本發明的目的是提供一種變速比大且具有較高扭矩分 配比，在單侧驅動輪在冰雪路面行進時越野能力大幅度增 強的具變動傳動比之防滑差速器。 為實現上述目的，本發明的技術解決方案是，一種具 變動傳動比之防滑差速器，通過行星齒輪與半軸齒輪間的 傳動比週期性的波動，使兩侧半軸錐齒輪上扭矩分配比例 成為行星錐齒輪轉角的週期性函數，來限制單側驅動輪的 5163-5619-PF2 6 1284719 滑轉’其主要結構包括差速器殼、其内設置有十字軸或一 字轴’變動傳動比的行星齒輪和半軸齒輪，其中，在嚙合 過程中行星齒輪和半軸齒輪傳動比以至少兩個周節為週期 波動變化，每一傳動比波動變化週期内的周節的個數應與 行星齒輪和半軸齒輪的齒數的公約數對應，每一個傳動比 皮動週期裏包含有一組參與工作的齒，該組的齒數與上述 周節的個數對應，每組中全部齒的合成工作範圍覆蓋整個 速比變化週期内行星輪和半軸輪的工作轉角範圍，每一齒 輪上各組齒的構成完全相同。 每組内每一個齒的工作轉角範圍可以在齒輪對的設計 過程中確定，相鄰齒的工作範圍應略有重疊。 本卷明的較佳方案是，上述周節的個數為3，因此， 行星齒輪和半轴齒輪的齒數均為3的倍數^ ^ 相鄰 一齒依一人為一個低齒’ 一個高齒和一個與前述低齒等高的 低齒，高齒與低齒之間為淺槽，兩個低齒之間為深槽；另 -種方案為-財的相鄰三齒依次為—個高齒，—個低齒 和-個與前述高齒等高的高#，高齒與低齒之間為深槽， 兩個高齒之間為淺槽。 值。半軸輪上齒的組數是杆s认 疋仃星輪個數的整倍數，每一行星 輪在相同轉角相位下工作。 行星輪上齒的組數為奇數，行星輪與—侧半軸齒輪傳 動比達到極大值時’與另—側半轴齒輪的傳動比達到極小 上述齒輪對的速比變化規律為·· 5163-5619-PF2 7 1284719 ^ · — 心 2 却(2) P — C.rarsin(z2iz)(2)/3) + C.(l —ra〇.sin(z2 ·〆2))] 式中A為半軸輪齒數，Z2為行星輪齒數，岬為半軸輪轉 #為行星輪轉角；半軸輪齒數A的取值範圍為g、12、 18’與之對應的行星輪齒數义的取值範圍為9、15; c 的取值範圍為〇·2〜〇·4; rat的取值範圍為〇 7〜1〇。 、變動傳動比錐齒輪齒廓的節線以下部分由解析曲線構 成，齒廓的節線以上部分是按要求的傳動比根據喷合原理 二才曰目對運動速度與相喃合的齒輪上的解析曲線齒廊的法線 向里垂直這一要求逐點求得的共軛齒廓曲線。與解析齒廓 共輛的齒廓曲線應滿足運動規律： ♦ = f[1 一C ·咖·sin(4(2) /3) + C · (1 -mi) ·sin(z2.卢(2))] 式中^為半軚輪齒數，％為行星輪齒數，,為半軸輪轉 角，〆2)為行星輪轉角；C的取值範圍為〇·2〜〇·4，rat的 取值範圍為0·7〜h0;半軸輪齒數A的取值範圍為9、12、 15 18 ,與之對應的行星輪齒數A的取值範圍為g、15。上 述解析曲線為直線、圓孤、_弧、漸開線、阿基米德和 對數螺旋線的組合。上述齒廓曲線均為凸齒面、適當的齒 頂厚度和重疊係數。由於一組參與工作的齒中每一對齒有 特疋的工作範圍，一組中的每一個齒有特定的齒廓形狀。 本發明的實現原理是將速比變化週期的周節的個數擴 大為兩個或兩個以上，從而使行星輪一轉中的速比變化次 數減少到傳統設計方法的二分之一或二分之一以下，於是 就可以在擴大速比變化範圍的同時大幅度降低了行星輪與 5163-5619-PF2 8 1284719 半軸輪的相對角加速度。 與現有技術相比，本發明的優點是· 該差速器為一種變動傳動比 在逆裔，其行星齒輪盥半 軸齒輪做變動傳動比傳動，傳 ” 闽 丄从抑+ J复化週期為兩個以上 周即。由於將速比週期改為兩個 Λ上周即，從而大幅度降 低了行星齒輪與半軸齒輪間的相對角加速度。 本發明的較佳實施例為一種r月 哭甘—田土 i 裡—周即變動傳動比差速 益，其灯星齒輪與半軸齒輪做變動傳 桃、 文勒得動比傳動，傳動比的 邊化週期為三個周節。由於將速^ ^ ^ ^ ^ ^ < ^双顯改為二個周節，從 而大幅度降低了行星齒輪盥车紅I _ 輪〃牛軸齒輪間的相對角加速度， 使兩半轴輪間的速比達到1 ·〗8 關u.85仍不發生齒面變尖的現 ^由於擴大了速比變化範圍’因而加高了對差速運動的 勢憂：料又w於職於大㈣分㈣㈣[輪㈣範圍 被加大了，也加寬了勢壘咸 、 了仃星輪因偶爾的振動而 越過勢曼的機率，握其了 财“了㈣的可靠性，因而大幅度提高 了差速器的鎖緊係數。 【實施方式】 下面將結合實施例及其附圖對本發明作進一步詳述： 實施例1 本實施例的具體結構請參見圖〗至圖3。本發明的且 變：傳動比之防滑差速器’包括差速器殼卜在差速器殼; 置於十子軸或一子轴5,行星齒輪4和半轴齒輪2，行 星齒輪4的背面與差速器殼j之間設置有球面塾片6,半 5163-5619-PF2 9 1284719 的背面與差速器之間設置有平墊片 半轴歯輪2構成若干對齒輪對。 3，上述行星 轴齒輪2 齒輪4和 齒數：：在本實施例中’選擇行星齒輪4和半軸齒輪2的 拥* f 、3的倍數，在嚙合過程中其傳動比以三個周節為 ’期波動變 卜 n > 抑 故母一個傳動比波動週期裏包含了三個相 W的每個齒都且 八有特定齒形的齒構成的一組參與工作的 齒。在同_ έΒ & 一 y 、、的二個齒中，由於每一個齒在速比變化週期 有特定的工作範圍，同一組中的齒高是變化的，齒形也 不相同。 ^ 同個齒輪上不同組間對應齒的齒形和齒高相同。行 生輪4上齒的組數為奇數，以保證當行星輪4與一侧半軸 :輪2的傳動比達到極大值時，與另-侧半軸輪2的傳動 達到極小值，這樣才能達到雨半轴輪之間最大的扭矩分 配比例。半軸輪2上齒的組數是行星輪4個數的整倍數， '母個行星輪4都在相同的轉角相位下工作，避免 各仃星輪4與半軸輪2之間的運動干涉。 具體在本實施例中，半軸齒輪2齒數的取值範圍為9、 12、15或18;行星齒輪4齒數的取值範圍為9、15;齒輪 對上對應的齒為三齒一組的齒形結構，三齒分別為一個高 齒和兩個等南的低齒，半軸齒輪2上高齒7與低齒8之間 為淺槽9，兩個低齒8之間為深槽1〇，行星齒輪4上高齒 13與低齒14之間為淺槽i 2，兩個低齒i 4之間為深槽11。 本實施例的實現原理是，將速比變化週期的擴大為三 個周節，從而使行星輪一轉中的速比變化次數減少到傳統 5163-5619-PF2 10 1284719 "又汁方法的三分之一，於是就可以在擴大速比變化範圍的 同時大te度降低了行星輪與半軸輪的相對角加速度，從而 達到本發明的目的。 本例中齒輪對的速比變化規律為： z2 却(2)二 K[1 一 C ·⑽· sin(z2〆2) /3) + C · (1 -. sin〇2 · 0(2))] 式中A為半軸齒輪齒數，义為行星齒輪齒數，〆”為半軸 輪轉角’多(2)為行星輪轉角。上述參數中c的取值範圍為〇 2 〜0·4 ; rat的取值範圍為〇·7〜1()時，左右兩側半轴輪 的速比變化範圍為0. 5〜2. 〇。 ^本發明的齒廓設計是基於上述傳動比關係來完成的。 定了齒輪對中一個齒輪的齒廊曲、線，就可以按給定的傳 ^ m 向上里垂直k 一要求逐點求得相應的共輕齒廊曲線。但必須 保δ且兩個齒面π at. ^ 面、適當的齒頂厚度和重疊係數。 由於一組參與工作的齒中 料I — 土 — ^母一對齒在速比變化週期中有Μ 疋的工作範圍，故三對齒 頁特 ^ ^ t齒中的母對齒廓有特定的形狀。本 .^ •齒廓節線以下部分為簡單的解柘 曲線。直線、圓弧和橢圓 解析 分是按要求的傳動比根據：=的組合曲線’節線《上部 合齒輪上的解析曲線齒靡==運動速度與相， 得的共軛齒廓曲線。 求、點求 說明本發明 而 依本發明作出的 非用於限定本發明： 具體試驗資料僅用以 5163-5619-PF2 11 1284719 〇·28 〜〇·3 18 15 0·18 〜〇·2 rat ' ——-— 速比變化範圍 鎖緊係數 〇·9 〜0·92 0.515 〜1.941 4. 5〜6. 9 〇· 86〜0. 88 0.538 〜1.857 3. 5〜4. 7 〇. 93 —----- 0.667〜1.500 2· 7〜3· 0 12 本實施例通過合理選擇行星輪與半軸輪齒數可使其具 有公約，3,因而將速比週期設計為3個周節個數。 本實施例中，利用兩側半轴輪之間週期性的傳動比波 動給差動運動造成勢壘，口女 ^ ，、有S兩侧半軸輪的扭矩差大於 傳動比的勢壘與摩擦力矩之知主 T刀矩之和時，差速器齒輪方能越過勢 壘連續差動運轉，否則口如少 μ & 貝h、此在一個傳動比波動週期即三個 周節内擺動。 實施例2 圖1、圖4和圖5為本發明實施例2的示意圖' 本實施例結構、工作焉理及+ 屌理及效果與上述實施例丨相 同，以在此不再贅述。 本實施例的具體結構請參見圖i、圖4、圖5。本發明 的具變動傳動比之防滑差速器’包括差速器殼i、在差速 器殼1内設置於十字軸或-字轴5,行星齒輪4和半軸齒 輪2’行星齒輪4的背面與差速器殼1之間設置有球面墊 片6’半轴齒輪2的背面與差速器之間設置有平墊片3,上 述行星齒輪4和半軸齒輪2構成若干對齒輪對。 具體在本實施例中’選擇行星齒輪4和半軸齒輪2的 齒數為3的倍數，在嚙合過程中其傳動比= ^ ~^固周即為週 期波動變化，故每一個傳動比波動週期襄人 ^ 3 了二個相鄰 5163-5619-PF2 12 1284719 特定齒形的齒構成的-組參與工作的齒。 特定的:作:ΓΓΓΓ 一個齒在速比變化週期中有 靶圍，.同一組中的齒高是變化的，齒形也不相 :-個齒輪上不同組間對應齒的齒形和齒高相同。行 的二拉* #』數為奇數，半抽輪上齒的組數是行星輪個數 的整倍數。 一本=施例與上述實施例的區別是，上述一組中的相鄰 一依人為個尚齒，一個低齒和一個與前述高齒等高的 门：I軸齒輪2上高齒21與低齒22之間為深槽24，兩 個同齒21之間為淺槽23，行星齒輪*上高齒託與低齒26 之間為木槽28 ’兩個高齒25之間為淺槽27。 本實施例的實現原理與上述相同，將速比變化週期的 擴大為三個周節，從而使行星輪4 一轉申的速比變化次數 減少到傳統設計方法的三分之一，於是就可以在擴大速比 蜒化範圍的同時大幅度降低了行星輪4與半軸輪2的相對 角加速度。 本實施例的其他結構特徵與實施例丨相同，原理及效 果相同，不再重複說明之。 上述實施例僅用於說明本發明，而非用於限定本發明。 【圖式簡單說明】 圖1是本發明的結構示意圖； 圖2是本發明齒輪對中半軸齒輪的結構示意圖； 圖3是本發明齒輪對中行星齒輪的結構示意圖。 5163—5619-PF2 13 1284719 圖4是本發明另一實施例的齒輪對中半軸齒輪的結構 示意圖；及 圖5是本發明另一實施例的齒輪對中行星齒輪的結構 示意圖。 【主要元件符號說明】 1〜差速器殼； 2〜半軸齒輪； 3〜平墊片； 4〜行星齒輪； 5〜十字軸或一字軸； 6〜球面墊片； 7、13、21、25〜高齒； 8、14、22、26〜低齒； 9、12、23、27〜淺槽； 10、11、24、28〜深槽。 5163-5619-PF2 14</ RTI> </ RTI> </ RTI> </ RTI> </ RTI> </ RTI> </ RTI> </ RTI> </ RTI> </ RTI> </ RTI> </ RTI> The invention relates to a wheeled vehicle and a non-slip differential for a non-slip differential device with a variable transmission ratio, specifically There are a variety of structures, such as internal friction type, transcendental type, electronic self-working principle can be combined, electronic brake anti-skid differential arguing and /, ABS brake system six factory and barrier type, each of them There are different shortcomings of the existence, such as the internal friction type of the two-seat type of the shackle, which can be subdivided into pre-loaded and non-reported, and the differentially-spotted site is cut into the work, which is expensive. The shortcomings of the latter, the price of the latter is higher, ^ test has the disadvantage of greater steering resistance, faster abortion wear; transcendence ι ... poor sin, unstable operation, complex structure; electronic automatic differential lock is also intent ^ The sub-defects are not stable enough and the structure is complicated; the brake anti-skid system that is in contact with the ABS brake system sisters has a large power consumption. This patent discloses a barrier-type variable ratio gear differential which has the advantages of a knot (four) 単, reliable performance and no negative traction, and the working principle is that the planetary gear and the half (four) wheel are cyclically shifted, and The speed change period on the planetary gears is an odd number of times, so when the transmission ratio of the planetary gears to the one side axle wheel reaches a maximum value, the transmission ratio with the other side half-pumping gear reaches a minimum value, thereby realizing the two side axles. Unequal torque distribution on the gears; the torque distribution on the side gears on both sides of the planetary gears will be reversed after half a speed ratio cycle. This periodic torque distribution change forms two 5163-5619-PF2 5 1284719 potential s. If the torque ratio on the two side axle wheels does not exceed the maximum torque distribution of the differential, the differential cannot be continuously poor. The movement limits the wheel slip. However, since the speed ratio change period of the variable transmission ratio differential currently used is one cycle, the relative acceleration and deceleration of the planetary wheel and the half shaft wheel are completed in each cycle, and the relative angular acceleration is large. The value is proportional to the range of the speed ratio change, and is proportional to the square of the number of fluctuations in the medium-speed ratio of the planetary gear. The relative angular acceleration of the car's father increases the relative curvature between the tooth faces, reduces the bearing capacity of the gear pair, and is prone to noise. Although the speed ratio fluctuation range can be improved to some extent by parameter optimization, the effect is limited, and further increasing the speed ratio range will rapidly increase the relative curvature between the tooth surfaces, and even generate ridge lines on the tooth surface. According to the traditional single-circle design method, the planetary gear is 7 teeth, the half-shaft gear is 12-tooth. The gear ratio of the differential gear is the maximum gear ratio produced by the differential motion between the two axle wheels. Only T:1•shape is 9 The 12-tooth differential of the tooth and half-shaft gear can only reach 1:1·31, which is still insufficient for off-road vehicles. SUMMARY OF THE INVENTION An object of the present invention is to provide a non-slip differential having a variable transmission ratio and a high torque distribution ratio, which is greatly improved in off-road capability when a single-side drive wheel travels on an ice or snow road. In order to achieve the above object, the technical solution of the present invention is a slip-resistant differential with a variable transmission ratio, and torque distribution on the side bevel gears on both sides by periodic fluctuations of the transmission ratio between the planetary gears and the half-shaft gears. The ratio becomes a periodic function of the angle of the planetary bevel gear to limit the 5163-5619-PF2 6 1284719 slip of the single-sided drive wheel. Its main structure includes the differential case, which is provided with a cross shaft or a word shaft. a planetary gear and a half-shaft gear, wherein the planetary gear and the half-shaft gear ratio fluctuate in a period of at least two cycles during the meshing process, and the number of the circumferential segments in each gear ratio fluctuation period should be The planetary gear and the half-shaft gear have a common divisor of the number of teeth. Each gear ratio has a set of teeth participating in the work. The number of teeth of the group corresponds to the number of the above-mentioned weeks, and the synthesis of all the teeth in each group The range covers the working angle range of the planetary gear and the half-shaft wheel in the entire speed ratio change period, and the composition of each group of teeth on each gear is identical. The working angle range of each tooth in each group can be determined during the design of the gear pair, and the working range of adjacent teeth should be slightly overlapped. The preferred embodiment of the present invention is that the number of the above-mentioned peripheral segments is three, and therefore, the number of teeth of the planetary gears and the half-shaft gears is a multiple of three ^ ^ adjacent one tooth is a low tooth of one person' a low tooth having the same height as the aforementioned low tooth, a shallow groove between the high tooth and the low tooth, and a deep groove between the two low teeth; another type of scheme is that the adjacent three teeth are in turn - a high tooth , a low tooth and a height # with the height of the aforementioned high tooth, a deep groove between the high tooth and the low tooth, and a shallow groove between the two high teeth. value. The number of sets of teeth on the half-axle is an integral multiple of the number of stars that the rod s recognizes, and each planet wheel operates at the same corner phase. The number of teeth on the planetary gear is an odd number. When the gear ratio of the planetary gear and the side axle gear reaches a maximum value, the transmission ratio of the gear to the other side gear is extremely small. The speed ratio of the gear pair is 5163- 5619-PF2 7 1284719 ^ · — heart 2 but (2) P — C.rarsin(z2iz)(2)/3) + C.(l —ra〇.sin(z2 ·〆2))] where A is The number of teeth of the half shaft, Z2 is the number of teeth of the planetary gear, 岬 is the half shaft rotation # is the planetary wheel angle; the value range of the number A of the half shaft gear is the range of the number of the planetary gear teeth corresponding to g, 12, 18' The value range of 9 and 15; c is 〇·2~〇·4; the value range of rat is 〇7~1〇. The lower part of the pitch line of the variable transmission ratio bevel gear is composed of an analytical curve, and the upper part of the pitch of the tooth profile is on the gear according to the required transmission ratio according to the principle of the spray-on. Analyze the normal to the vertical of the curved tooth gallery. This requires a conjugate tooth profile obtained point by point. The tooth profile curve shared with the analytical tooth profile should satisfy the motion law: ♦ = f[1 - C · coffee · sin(4(2) /3) + C · (1 -mi) · sin(z2. Lu (2 ))] where ^ is the number of semi-turned teeth, % is the number of planetary gear teeth, is the half-axis wheel angle, 〆 2) is the planetary wheel angle; C is in the range of 〇·2~〇·4, The value range is 0·7~h0; the number of teeth of the half-axis gear is in the range of 9, 12, 15 18 , and the number of teeth A corresponding to the planetary gears is in the range of g and 15. The above analytical curves are a combination of straight lines, rounded orphans, _arcs, involutes, Archimedes, and logarithmic spirals. The above tooth profiles are convex tooth faces, appropriate crown thicknesses and overlap coefficients. Since each pair of teeth participating in the work has a characteristic working range, each tooth in the set has a specific profile shape. The implementation principle of the present invention is to expand the number of circumferential sections of the speed ratio change period to two or more, thereby reducing the number of speed ratio changes in one revolution of the planetary gear to one-half or two of the conventional design method. In less than one part, the relative angular acceleration of the planetary gear and the 5163-5619-PF2 8 1284719 half-axle can be greatly reduced while expanding the range of the speed ratio. Compared with the prior art, the advantage of the present invention is that the differential is a variable transmission ratio in the reverse genus, and the planetary gear 盥 half shaft gear is changed in the transmission ratio transmission, and the 复 闽丄 抑 + J recombination cycle is More than two weeks, since the speed ratio period is changed to two Λ last week, the relative angular acceleration between the planetary gear and the side gear is greatly reduced. A preferred embodiment of the present invention is a r month crying - Tiandi i - Week is the change of the transmission ratio difference speed, the lamp star gear and the half shaft gear change the transmission peach, Wenle to the specific transmission, the transmission cycle of the transmission cycle is three weeks. Because the speed will be ^ ^ ^ ^ ^ ^ < ^ Double display changed to two quarters, which greatly reduced the relative angular acceleration between the planetary gears, the red I _ rim shaft gears, and the speed ratio between the two axle wheels 1 ·〗 8 u u.85 still does not occur in the tooth surface sharpening ^ due to the expansion of the speed ratio change range 'and thus increased the potential for differential movement: expected to work in the big (four) points (four) (four) [round (4) The scope has been enlarged, and the barrier has been widened. The comet has crossed the occasional vibration. Mann's chances, shook his fortune "(iv) the reliability, thus greatly improving the coefficient locking differential. [Embodiment] The present invention will be further described in detail below with reference to the embodiments and the accompanying drawings. Embodiment 1 For the specific structure of this embodiment, please refer to FIG. 3 to FIG. The invention also has the following advantages: a transmission ratio anti-skid differential 'includes a differential case in the differential case; is placed on a ten-axis or a sub-axis 5, a planetary gear 4 and a side gear 2, and a planetary gear 4 A spherical cymbal 6 is disposed between the back surface and the differential case j, and a flat washer half shaft 歯 wheel 2 is disposed between the back surface of the half 5163-5619-PF2 9 1284719 and the differential to form a pair of gear pairs. 3. The above-mentioned planetary gear 2 gear 4 and the number of teeth: In the present embodiment, 'the multiples of the planetary gear 4 and the side gear 2 are selected *f, 3, and the transmission ratio is three weeks in the meshing process. 'Period fluctuations n > In the case of a gear ratio fluctuation period, a set of three teeth of each phase W and eight teeth with a specific tooth shape constitute a group of working teeth. In the two teeth of the same _ έΒ & y , , since each tooth has a specific working range in the speed ratio change period, the tooth height in the same group is changed, and the tooth shape is also different. ^ The tooth shape and tooth height of the corresponding teeth between different groups on the same gear are the same. The number of teeth on the row wheel 4 is an odd number to ensure that when the transmission ratio of the planetary gear 4 and the one side axle: the wheel 2 reaches a maximum value, the transmission with the other side wheel axle 2 reaches a minimum value, so that The maximum torque distribution ratio between the rain axles is reached. The number of sets of teeth on the half-shaft wheel 2 is an integral multiple of the number of planetary gears 4, and the 'mother planetary gears 4 are all operated at the same corner angle to avoid motion interference between the respective star wheel 4 and the half-shaft wheel 2 . Specifically, in the embodiment, the number of teeth of the side gear 2 ranges from 9, 12, 15 or 18; the number of teeth of the planetary gear 4 ranges from 9, 15; the corresponding tooth on the pair of gears is a set of three teeth. The tooth-shaped structure, the three teeth are respectively a high tooth and two low teeth of the south, and the shallow groove 9 between the high tooth 7 and the low tooth 8 on the side gear 2, and the deep groove 1 between the two low teeth 8 〇, between the high teeth 13 and the low teeth 14 on the planetary gear 4 is a shallow groove i 2 , and between the two low teeth i 4 is a deep groove 11 . The implementation principle of this embodiment is to expand the speed ratio change period to three quarters, so that the number of speed ratio changes in one revolution of the planetary gear is reduced to the conventional 5163-5619-PF2 10 1284719 " In one part, the relative angular acceleration of the planetary wheel and the half-shaft wheel can be reduced by a large degree while the range of the speed ratio is increased, thereby achieving the object of the present invention. In this example, the speed ratio of the gear pair changes: z2 but (2) two K[1 - C · (10) · sin(z2〆2) /3) + C · (1 -. sin〇2 · 0(2) In the formula, A is the number of teeth of the half shaft gear, meaning the number of teeth of the planetary gear, 〆" is the angle of the half shaft wheel. (2) is the angle of the planetary wheel. The value of c in the above parameters ranges from 〇2 to 0·4; When the value range of rat is 〇·7~1(), the ratio of the speed ratio of the left and right side axle wheels is 0. 5~2. 〇. ^ The tooth profile design of the present invention is based on the above transmission ratio relationship. When the gears and lines of one gear of the gear pair are fixed, the corresponding common light tooth corridor curve can be obtained point by point according to the requirement of the vertical transmission k. Tooth surface π at. ^ surface, appropriate tip thickness and overlap coefficient. Because a group of teeth involved in the work I - soil - ^ mother pair of teeth in the speed ratio change period has a working range of Μ 故, so three The parent-to-tooth profile in the tooth surface has a specific shape. This is a simple solution curve for the following parts of the tooth profile line. The straight line, arc and ellipse resolution are the required gear ratios. root The combination curve of := 'the pitch line' is the analytical curve of the upper part of the gear 靡 == the speed of movement and the phase, the obtained conjugate tooth profile curve. The purpose of the invention is not limited by the invention. The invention: the specific test data is only used for 5163-5619-PF2 11 1284719 〇·28 〇·3 18 15 0·18 〇 2 2 rat ' ——-- speed ratio change range locking coefficient 〇·9 〜 0·92 0.515 ～1.941 4. 5~6. 9 〇· 86~0. 88 0.538 〜1.857 3. 5~4. 7 〇. 93 —----- 0.667~1.500 2· 7~3· 0 12 In this embodiment, the number of teeth of the planetary wheel and the half shaft can be selected to have a convention, 3, and thus the speed ratio period is designed to be 3 weeks. In this embodiment, the periodicity between the two side axles is utilized. The gear ratio fluctuation causes a potential barrier for the differential motion, and the torque difference between the two side axles of the S side is greater than the sum of the barrier of the gear ratio and the T-knife of the friction torque, the differential gear It can continue to run differentially across the barrier, otherwise the mouth is as small as μ & h, this swings in a gear ratio fluctuation period, that is, three weeks. 2, FIG. 4 and FIG. 5 are schematic diagrams of the second embodiment of the present invention. The structure, the working process, the + processing and the effect of the present embodiment are the same as those of the above embodiment, and will not be further described herein. For details, please refer to FIG. i, FIG. 4 and FIG. 5. The slip-resistant differential with variable gear ratio of the present invention includes a differential case i, which is disposed on the cross shaft or the --axis 5 in the differential case 1. a spherical washer is disposed between the back surface of the planetary gear 4 and the side gear 2' planetary gear 4 and the differential case 1; a flat washer 3 is disposed between the rear surface of the side gear 2 and the differential, and the above-mentioned planet The gear 4 and the side gear 2 form a pair of pairs of gears. Specifically, in the present embodiment, 'the number of teeth of the planetary gear 4 and the side gear 2 is selected to be a multiple of 3. In the meshing process, the gear ratio = ^ ~ ^ solid cycle is the cyclic fluctuation, so each gear ratio fluctuation period 襄The person ^ 3 has two adjacent 5163-5619-PF2 12 1284719 teeth of a particular tooth shape - the group of teeth involved in the work. Specific: for: ΓΓΓΓ One tooth has a target circumference in the speed ratio change period. The tooth height in the same group is changed, and the tooth shape is not in phase: the tooth shape and tooth height of the corresponding tooth between different groups on one gear the same. The number of rows of the two pulls is an odd number, and the number of teeth on the half-pump is an integral multiple of the number of planet wheels. A difference between the embodiment and the above embodiment is that the adjacent one of the above groups is a tooth, a low tooth and a door having the same height as the high tooth: the high tooth 21 on the I-axis gear 2 There is a deep groove 24 between the low teeth 22, a shallow groove 23 between the two identical teeth 21, and a wooden groove 28 between the high tooth support and the low tooth 26 on the planetary gear * 'the shallow groove between the two high teeth 25 27. The implementation principle of this embodiment is the same as the above, and the speed ratio change period is expanded to three weeks, so that the number of revolutions of the planetary gear 4 is reduced to one third of the conventional design method, so that The relative angular acceleration of the planetary gear 4 and the axle wheel 2 is greatly reduced while the speed ratio is increased. Other structural features of the embodiment are the same as those of the embodiment, and the principles and effects are the same, and the description thereof will not be repeated. The above examples are merely illustrative of the invention and are not intended to limit the invention. BRIEF DESCRIPTION OF THE DRAWINGS Fig. 1 is a schematic view showing the structure of a gear centering half shaft gear of the present invention; Fig. 3 is a structural schematic view of a gear centering planetary gear of the present invention. 5163-5619-PF2 13 1284719 FIG. 4 is a schematic view showing the structure of a gear centering half shaft gear according to another embodiment of the present invention; and FIG. 5 is a schematic view showing the structure of a gear centering planetary gear according to another embodiment of the present invention. [Main component symbol description] 1~ differential case; 2~ half shaft gear; 3~ flat washer; 4~ planetary gear; 5~ cross shaft or slot; 6~ spherical washer; 7, 13, 21 25, high teeth; 8, 14, 22, 26 ~ low teeth; 9, 12, 23, 27 ~ shallow grooves; 10, 11, 24, 28 ~ deep grooves. 5163-5619-PF2 14